COMPARATIVE ANALYSES OF TWO IMPROVED CO 2 COMBINED COOLING, HEATING, AND POWER SYSTEMS DRIVEN BY SOLAR ENERGY

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1 S93 Introduction COMPARATIVE ANALYSES OF TWO IMPROVED CO 2 COMBINED COOLING, HEATING, AND POWER SYSTEMS DRIVEN BY SOLAR ENERGY by Wanjin BAI a* and Xiaoxiao XU b a School of Mechanical and Vehicle Engineering, Chongqing Jiaotong University, Chongqing, China b Key Laboratory of Low-Grade Energy Utilization Technologies and Systems, Chongqing University, Chongqing, China Original scientific paper To make use of solar energy fully and efficiently, the two improved combined cooling, heating, and power systems (CCHP) are proposed by adding a gas heater and an extraction turbine, based on the transcritical CO 2 ejector refrigeration system. A relatively high pressure fluid is extracted by the extraction turbine as a primary stream of ejector to improve the ejector performance. In the meantime, the gas heater absorbs low temperature exhaust heat to increase the extraction turbine s output work. Comparative studies on the thermal efficiency and exergy efficiency of the two improved systems show they are more efficient alternatives for the transcritical CO 2 ejector refrigeration system. The CCHP-B system has relatively broad working condition, higher thermal efficiency and exergy efficiency than that of CCHP-A system. Key words: CCHP, CO 2, supercritical, exergy efficiency, thermal efficiency In order to solve environmental and energy problems, as a kind of clean renewable energy, solar energy has a rapid progress in research and development. The combined cooling, heating, and power (CCHP) system driven by solar energy with higher system efficiency, more eco-friendly environment, and better economic feasibility, offers an effective way to realize sustainable development [, 2]. More and more literature studies can be found using the natural working fluid CO 2 in the Rankine cycle or the ejector refrigeration cycle [3]. The lower critical temperature (3. o C) of CO 2 results in the transcritical or supercritical Rankine cycle and the transcritical ejector refrigeration cycle or in supercritical cogeneration system. Kim et al. [4] presents the optimization processes for supercritical CO 2 Rankine cycles. Nami et al. [5] proposes and analyzes a novel co-generation system, including a gas turbine, a heat recovery steam generator and a supercritical CO 2 cycle (GT-HRSG/SCO 2 ). Padilla et al. [] perform a comprehensive thermodynamic analysis and a multi-objective optimization to study the proposed supercritical CO 2 Brayton cycles integrated with an ejector. The transcritical CO 2 ejector system has been widely used in air conditioning in recent years. But, the major difference between the transcritical CO 2 cycle and conventional refrigerant cycle is that up to 80 o C ~50 o C is reached at the outlet temperature of CO 2 compressor with its specific thermal * Corresponding author, baiwking@3.com

2 S94 properties near the pseudo-critical point. This system ignores the fact that the exit heat from the compressor is cooled by the gas cooler, which means the heat would go to waste. Using lowgrade heat sources to further raise the heat temperature of the compressor outlet is an attractive alternative solution, and the heat grade is enhanced for producing heating output, cooling output and power output simultaneously. Xu et al. [7] proposed two improved CCHP systems by adding a gas heater and an extraction turbine. The extraction turbine is to extract a high pressure stream as a primary fluid of ejector which is benefit to obtain large refrigeration output, and continues to expand the remaining fluid of the extraction turbine to a lower pressure for more power output. In addition, the supercritical CO 2 stream from the compressor outlet is further heated by the gas heater, which upgrades the heat quality. Our improved CCHP systems using supercritical CO 2 have owned the authorized patent [8]. Based the previous study on parametric analysis and exergy analysis for one of the improved CCHP systems [7], this paper attempts to compare the two improved CCHP using supercritical CO 2 to investigate the influence of the thermodynamic parameters on the performance and exergy destruction of the CCHP systems. Cycle description Figures and 2 show the schematic diagrams of the improved CCHP cycles, figs. (a) and 2(a), and the processes of the corresponding cycles in a temperature-specific entropy diagrams, figs. (b) and 2(b). As can be seen in figs. (a) and 2(a), the major difference between the two CCHP cycles is that the extracted fluid from the extracted turbine with the fluid exiting from the evaporator are fed to ejector in CCHP-A cycle, fig., whereas the fluid exiting from the extracted turbine with the saturated vapor from the gas-liquid separator are fed to compressor directly in CCHP-B cycle, fig. 2. Thermodynamics analysis Ejector model and system model The mathematical models for ejector and overall system are described and presented in authors former paper [2]. The principle of the ejector model is introduced by Keenan et al. [9]. Three independent efficiencies are assumed as 0.9, 0.85, and 0.85 for the motive nozzle, mixing section, and diffuser, respectively. Solarenergy T 2 4a Turbine 3a 2 3b 4b 3c 4c Heater Gas cooler Heating user Auxiliary heater 9 8b Qg Gas heater Ejector Heat storage Compressor Liquid-vapor separator 2 3 Metering Metering valve valve 8a c 8a 0 3b (2, 3) 8b 9 4b 4c 3a 4a Q h Qgc (a) Evaporator Cooling user Q e (b) Figure. The CCHP-A cycle; (a) schematic diagram, (b) temperature-entropy chart S

3 S95 Solarenergy T 2 4a Turbine 3a Auxiliary heater 2 Q g Gasheater Heat storage Compressor 3a 4a Heater Heatinguser Gascooler Q h 3b 4b 3c Q gc 4c 9 8b Ejector Liquid-vapor separator 2 3 Metering Metering valve valve 8a a 3c 3b 4b 4c (3) 2 0 8b (a) Evaporator Coolinguser Figure 2. The CCHP-B cycle, (a) schematic diagram, (b) temperature-entropy chart System stability conditions Q e (b) Li and Groll [0] points out that the mass conservation constraint x = /( + µ) is used to control the stability of the transcritical CO 2 ejector refrigeration system. In this work, we introduce the gas heater and the extraction turbine into the Li s modified cycle. The equation x > /( + µ) is equivalent in function to m 3 > 0, that is, the mass-flow rate of the bypass is larger than zero. CCHP-A cycle m = e( + ) x m () [ ] 3 5 The stable operation condition is written according to Li and Groll [0] (m 3 > 0): [ µ ] e( + ) x m > 0 (2) The ejector outlet quality must satisfy eq. (3) in order to maintain the system stability and realize the cycle: x > e ( + µ ) (3) Since the ejector outlet quality is related to the extraction ratio, it demonstrates that the extraction ratio of CCHP-A cycle can not be too low otherwise it would be unable to satisfy the eq. (3). CCHP-B cycle m = em + x + µ x (4) ( ) 3 The stable operation condition is written according to Li and Groll.[0] (m 3 > 0): + x + µ x > 0 (5) The ejector outlet quality must satisfy eq. () in order to maintain the system stability and realize the cycle: x > + µ () Since the ejector outlet quality is not related to the extraction ratio, the CCHP-B cycle has relatively broad range and application compared to CCHP-A cycle. S

4 S9 System performance Considering the improved CCHP cycles combined with the vapor compression cycle and the power cycle, is defined separately by the refrigeration and the energy utilization efficiency, η EUE, are used to characterize the thermal efficiency. The ratio of, ξ, between the improved CCHP cycle and the ejector cycle is defined to indicate the system performance improvement: Qh + Wt Wc ηeue = (7) Q g ξ = (8) EJE The exergy efficiency, η EXG, is a criterion for the system performance: η E + E + W W Q,e Q,h t c EXG = (9) EQ,g E Q,i T = Qi T 0 Q,i ( ) ( ) (0) Ei = mi hi h0 T0 si s0 () I = T Q + E E W Results and discussion (2) 0 i i in,i out,i i i i i T Q,i Comparison between the two improved CCHP cycles The working conditions of the two CCHP cycles shown in tab. are reasonably assumed to maintain the cycles run stably. Table. The simulation conditions of the improved CCHP cycles Parameters Amount (Type-A cycle) Amount (Type-B cycle) Mass-flow rate of working fluid.4 kg/s.4 kg/s Environment temperature 5 5 Environment pressure 0.03 MPa 0.03 MPa Turbine inlet temperature Turbine inlet pressure 2 MPa 2 MPa Turbine extraction pressure 8.2 MPa 8.4 MPa Turbine extraction rate Ejector inlet temperature 3 3 Ejector back pressure 4.4 MPa 4.MPa Heater outlet temperature Evaporation temperature 5 5 Turbine isentropic efficiency Compressor isentropic efficiency Approach temperature difference of heat exchanger 0 0 Mass-flow rate of flue 5 kg/s 5 kg/s

5 S97 Table 2 shows the performance of the improved CCHP cycles. As the turbine extraction ratio of CCHP-A cycle is higher than that of CCHP-B cycle, more high temperature refrigerant extracting from the turbine can supply more heating output in the CCHP-A cycle, whereas more refrigerant working in turbine results in more turbine output in the CCHP-B cycle. Although the difference of the refrigerant output in two cycles is very little, the and the exergy efficiency of CCHP-B cycle are higher than that of the CCHP -A cycle. Table 2. The performance of the improved CCHP cycles Parameters Amount (CCHP-A cycle) Amount (CCHP-B cycle) Turbine power kw kw Refrigerating output 7.44 kw 74.5 kw Heating output kw kw Heat absorption from heat source kw 28.2 kw Compressor power.7 kw 7.82 kw Energy utilization efficiency Exergy efficiency Figure 3 shows the influence of extraction rate on the both system performance. It notes that the ejector outlet quality of CCHP-A cycle and CCHP-B cycle should satisfy the condition listed in eq. (3) and eq. (), respectively. The ejector outlet quality of CCHP-A cycle cannot be less than 0.8. As the extraction rate increases, the refrigeration output increase due to extracting more supercritical fluid CO 2 from turbine as the driving force of the ejector which entrained more stream from the evaporator. But, more supercritical fluid CO 2 extracted decrease the turbine power output, which far outpaces the growth rate of the refrigeration output. Thus, the decreases with the increasing of extraction rate. Compared to the CCHP-A cycle, the CCHP-B cycle is more suitability for the high demand of the refrigeration output. Efficiency Q e Q h W c W t η EXG η EUE Q e,q h,w c,w t [kw] (a) Extraction ratioofturbine (b) Extraction ratio of turbine Efficiency Figure 3. Effect of extraction rate on the system performance; (a) CCHP-A cycle (b) CCHP-B cycle Based on the exergy destruction of components in the systems as shown in tab. 3, the irreversibility of heat transfer is the major source of exergy destruction in both systems. The exergy destruction ratio of three heat exchangers in CCHP-B system is 7.% and that of CCHP-A system is 9.88%. The results reveal that the system exergy destruction could be reduced by decrease the heat transfer temperature difference in three heat exchangers. The most Q e Q h W c W t η EXG η EUE Q e,q h,w c,w t [kw]

6 S98 significant difference of the two systems lies in the exergy destruction of ejector outperformed by 4.% than those of turbine in CCHP-A system, whereas, lowered by.5% in CCHP-B system. Ma et al. [] points out that the exergy destruction of the turbine is lower than that of the ejector under general operating conditions. But, the result presents in CCHP-B system shows just the reverse. This could be explained by more working fluid CO 2 expanded in the turbine further increases the net power output, which also leads to the increase of the exergy efficiency. Table 3. The exergy destruction of the two improved CCHP systems Components CCHP-A cycle exergy destruction, [kj] CCHP-A cycle exergy destruction ratio, [%] CCHP-B cycle exergy destruction, [kj] CCHP-B cycle exergy destruction, [%] Compressor Ejector Evaporator Gas cooler Gas heater Heater Turbine Throttle Total Comparisons between the improved CCHP systems and the transcritical CO 2 ejector system The previous two improved CCHP systems in this paper are based on the transcritical CO 2 ejector system proposed Li and Groll [0]. It is necessary to compare the improved CCHP systems with the transcritical CO 2 ejector system so as to estimate whether the expander will improve the system performance. The calculation is based on the same conditions including the ambient temperature 5, the expander isentropic efficiency 0.85, the compressor isentropic efficiency 0.8, and the evaporating temperature 5 and the gas cooler outlet temperature 3. The of the combined cycle is: Qe AB / = (3) Wc Wt The of the conventional injection refrigeration cycle is: Q e conventional = (4) Wc The ratio of CCHP-A and CCHP-B cycle to conventional transcritical ejector refrigeration cycle is expressed, respectively: Ratio / A = A conventional (5) Ratio = / () B B conventional Figure 4 shows that there is optimum high-side pressure and maximum in the transcritical CO 2 ejector system. The maximizing high-side pressures of the improved CCHP systems are higher than that of the transcritical CO 2 ejector system. This is because the network equals the difference of compressor work and turbine power output, not equals compressor work. For CCHP-A cycle, the optimum high-side pressure is 4 MPa, for CCHP-B cycle is 2 MPa. Under a given condition in this paper, B is the largest among three cycles.

7 S99 It also can be seen that the of the improved CCHP systems are higher than that of the transcritical CO 2 ejector system with increase highside pressure. The difference between the improved CCHP system and transcritical CO 2 ejector system becomes more and more large as the high-side pressure increases. The Ratio A increases from and Ratio B from To be mentioned that, the improved CCHP system has evident advantages over the transcritical CO 2 ejector system when they run under the conditions of the high-side pressure B A Ratio B Ratio A High-side pressure (MPa) Figure 4. Comparison chart of the Ratio Conclusions y The two improved CCHP system using supercritical CO 2 are proposed. The CCHP-B system has higher thermal efficiency and exergy efficiency than that of CCHP-A system. y The stability conditions of the two improved CCHP cycle are obtained. The CCHP-B system has relatively broad working condition than that of CCHP-A system. y The improved CCHP systems exist optimal high-side pressures that give maximum, and the maximizing high-side pressure of the improved CCHP systems are higher than that of the transcritical CO 2 ejector system. y Both extraction pressure and higher extraction rate are helpful to gain more refrigeration in the improved CCHP systems. Acknowledgment The authors gratefully acknowledge National Natural Science Foundation of China (No ), China Postdoctoral Science Foundation (No. 207M0590) and Chongqing Postdoctoral Scientific Research Foundation (Project No. Xm207045) for providing the financial support for this study. Nomenclature E exergy, [kw] e extraction ratio, [ ] h specific enthalpy, [kjkg ] m mass-flow rate, [kgs ] p pressure, [MPa] Q heat load, [kw] s specific entropy, [kjkg K ] T temperature, [ o C] W power, [kw] Greek symbols η efficiency µ entrainment ratio ξ ratio Subscripts c compressor e evaporator g gas heater h heater Q heating source t turbine u velocity, [ms ] x vapor quality, [ ] Abbreviations coefficient of performance, [ ] EJE ejector cycle EUE energy utilization

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